The present invention comprises improvements to generally known two-stage adjustable length connecting rods in which the adjustable element is an eccentric connection of the rod end to the piston pin that is actuated by the periodically reversing gas and inertial dynamic forces acting on the rod and eccentric assembly, and latched in either of two positions resulting in two different effective rod lengths and corresponding compression ratios. U.S. Pat. No. 7,685,974 issued to Berger provides extensive background information including prior art, utility of two-stage variable compression ratio for improved engine efficiency, and nature of the periodic gas and inertial dynamic forces operating on the rod and eccentric assembly. '974 is incorporated by reference, and a brief summary follows. The discussion is directed to spark ignition throttle controlled internal combustion engines typically used in light duty vehicles, although it is obvious that the invention is applicable to compression ignition engines and other reciprocating machines such as compressors.
Internal combustion engine compression ratio is defined as the volume between the piston and the cylinder head at bottom dead center divided by the volume at top dead center. In typical prior art engines the compression ratio is fixed by the engine design and is not an adjustable operating variable. In high compression engines the volume at top dead center is relatively small, and the compressed fuel-air mixture is at a higher temperature and pressure than in low compression engines at the moment of ignition. Increasing temperature and pressure increase the engine thermodynamic efficiency by increasing the combustion pressure that drives the piston toward bottom dead center on the power stroke, thereby producing more useful work from a given quantity of fuel. The temperature and pressure must, however, remain below the knock-limit. Below the knock limit, the fuel-air mixture combusts as a progressive flame front passing through the combustion chamber volume, and results in a smooth pressure rise to drive the piston. Above the knock limit, a portion of the fuel-air mixture in the combustion chamber undergoes very rapid bulk combustion that produces a pressure spike resulting in audible noise and potential engine damage. The knock-limit is affected by a number of variables other than compression ratio including fuel chemical composition, inlet manifold pressure, engine temperature, and ignition timing relative to piston position. Modern engines typically have knock sensors that detect incipient knock. These allow the engine control module to automatically make adjustments, e.g. varying ignition timing, to keep the engine operating at the highest possible efficiency without knock over the full engine load and speed range, while compensating for other variables such as fuel chemistry and engine temperature. Late or retarded ignition timing from the maximum torque value reduces knock, but also reduces engine efficiency and increases heat transfer losses.
The fixed compression ratio of prior art engines is a compromise that balances maximum output torque against light load efficiency. Maximum output torque at a given speed occurs when the throttle is fully open, thereby providing the maximum gas pressure in the inlet system. This gas may be a fuel-air mixture in manifold or port injection systems, or air alone in direct cylinder injection systems in which fuel is added to the air in the cylinder. In a naturally aspirated engine this pressure is slightly below atmospheric pressure because of flow losses, and in a turbocharger or supercharger pressure boosted engine it is above atmospheric pressure. The cylinder at the beginning of the compression stroke is filled with a charge of gas slightly below the inlet manifold pressure, since inlet system flow losses prevent complete pressure equilibration between the inlet system and the cylinder. The compression ratio is chosen so that at the end of the compression stroke the gas charge is at a temperature and pressure that results in maximum output torque with optimized ignition timing and without knock. A pressure boosted engine generally requires a lower compression ratio to avoid knock at maximum torque since the initial gas pressure and charge mass is higher. At light load the throttle reduces the inlet system pressure and the initial cylinder pressure resulting in a smaller charge mass. This leads to lower than optimum gas charge temperature and pressure at the end of the compression stroke, and consequent reduced efficiency. Since the typical duty cycle of light duty vehicles consists of predominantly light load operation, the non-optimum compression ratio results in significant efficiency losses.
It is generally known that that a capability to adjust compression ratio in a running engine as a function of operating conditions has potential to improve both light load efficiency and maximum output torque compared to a fixed compression ratio. High compression at light loads compresses the low mass light load gas charge into a reduced combustion chamber volume. This increases the gas charge temperature and pressure at the end of the compression stroke to a more optimum level and increases the thermodynamic efficiency. Reduced compression at maximum output torque compresses the maximum load gas charge into an increased combustion chamber volume. This allows a larger charge mass without exceeding the optimum temperature and pressure at the end of the compression stroke, allowing higher output torque. This is a particular advantage in pressure boosted engines since it permits a smaller displacement engine to provide the required maximum output torque, offering potential engine size and weight savings.
Various approaches to adjusting the compression ratio are known, and are summarized in '974. Two-stage adjustable length connecting rods are the subject of '974 and a number of other patents, and are attractive because they require minimum changes to the engine and add little overall size or weight. This approach employs two-stage adjustment in which the rod may be switched between two fixed lengths, since it is simpler than continuous adjustment and provides nearly the same benefit. As in a conventional fixed compression ratio engine, the crankshaft axis is fixed relative to the cylinder head. An increase in connecting rod length reduces the volume at top dead center and increases the compression ratio, while a decrease in length increases the volume at top dead center and reduces the compression ratio. Many mechanisms are disclosed in the prior art to carry out the adjustment, and a number use an eccentric bushing between the piston wristpin and the small end of the connecting rod that is rotated relative to the connecting rod in a journal bearing to adjust the effective connecting rod length. Although there are exceptions, typical prior art mechanisms are self-powered in that the alternating compressive and tensile forces on the connecting rod and eccentric bushing during the engine cycle generate torque that rotates the eccentric to change the effective rod length. Mechanical stops are incorporated to limit the eccentric bushing rotation to a specific angle. A bi-stable mechanical latch mounted on the moving connecting rod locks the eccentric into either the high compression or the low compression position, and is reset by a stationary trigger mechanism that interacts with the latch on the moving connecting rod to change the compression ratio. When the latch is reset it is biased to disengage the eccentric, and then engage it again after the alternating rod forces rotate the eccentric to the new selected position. The latch typically disengages only under low load as the alternating rod load reverses, and the eccentric will only move toward the new selected position when the alternating rod load is in the correct direction. Compressive rod force is required to rotate the eccentric within the journal bearing to reduce the effective rod length, and tensile rod force is required to rotate the eccentric in the opposite direction to increase the effective rod length. The latch reengages the eccentric only when the eccentric reaches the new position. One or more engine revolutions after latch reset are required to obtain the connecting rod force variations needed to release the latch and rotate the eccentric to the new position so that the latch may reengage.
Three principal approaches are proposed in the prior art to allow a stationary trigger mechanism to reset the latch on the moving connecting rod to change the compression ratio: hydrostatic fluid interaction, hydrodynamic fluid interaction, and mechanical cam interaction. Hydrostatic fluid interaction employs passages in the engine block, crankshaft and connecting rod such that controllable oil pressure may be transmitted from a stationary source through the rotating interfaces to a hydraulic plunger that shifts the latch mechanism between positions on the moving connecting rod. Hydrodynamic fluid interaction employs stationary oil nozzles mounted to the engine block that direct momentary oil jets to exert forces on control surfaces carried by the moving connecting rod that move to shift the latch between positions. Mechanical cam interaction employs cam surfaces mounted to the engine block that may be controllably shifted to contact and shift the latch between positions on the moving connecting rod. Hybrid approaches, e.g. mechanical cam surfaces that shift valves in the connecting rod that redirect pressurized oil to hydrostatically shift the latches between positions, are also known.
The two-stage adjustable length connecting rods of '974 and related prior art have an inherent indeterminacy related to the fact that alternating compressive and tensile forces on the connecting rod and eccentric bushing during the engine cycle are used to rotate the eccentric to change the effective rod length. Ideally, one or two engine revolutions after latch reset are required to obtain the connecting rod force variations needed to release the latch and rotate the eccentric to the new position so that the latch may reengage. If, however, the eccentric does not rotate enough to reach the new position and engage the latch while the rod force is in the correct direction, the subsequent force reversal will reverse the initial eccentric rotation direction. This causes a loss of the desired rotation, and may result in situations in which the eccentric rotational position cycles with each rod force direction, and full rotation and latch engagement is either slow or never completed. Such unproductive cycling may also cause wear and noise. Since maximum compressive rod forces include compression and power stroke gas loading and are typically much larger than primarily inertial tensile rod forces, it may be more difficult to lengthen the rod and increase the compression ratio.
The relationship between the eccentric geometry and the coefficient of friction in the journal bearing between the eccentric and the connecting rod is critical in achieving reliable eccentric rotation while minimizing stress on the latch. The parameters are shown in FIG. 12:
R is the outside radius of the eccentric 108 which rotates within the journal formed by the small end 109 of the connecting rod body 110;
r is the eccentric offset between the journal center of the piston wrist pin 107 and the center of the outside diameter of the eccentric 108;
α is the rotational angle between a line 1200 perpendicular to the centerline 1201 of the connecting rod body 110;
F is the instantaneous value of the tensile or compressive force on the connecting rod 100;
μ is the coefficient of friction between the eccentric 108 rotating within the small end 109 of the connecting rod body 110;
Mf is the torque on the eccentric 108 generated by the friction between the eccentric 108 and the small end 109 of the connecting rod body 110; and
Mr is the reaction torque on the eccentric 108 generated by the force F acting on the eccentric offset.Mf=FμR Mr=Fr cos α
If the reaction torque Mr is greater than the friction torque Mf the eccentric will rotate in the direction of the reaction torque, and if it is less it will not rotate. It is useful to define a parameter Z as the ratio of the reaction torque to the friction torque:
  Z  =                    M        r                    M        f              =                            Fr          ⁢                                          ⁢          cos          ⁢                                          ⁢          α                          F          ⁢                                          ⁢          μ          ⁢                                          ⁢          R                    =                        r          ⁢                                          ⁢          cos          ⁢                                          ⁢          α                          μ          ⁢                                          ⁢          R                    
If Z is greater than 1 rotation will take place, and if it is less than 1 it will not. A large Z therefore provides higher assurance of successful rotation between latched positions. A smaller Z however, counteracts a larger portion of the reaction torque with the friction torque and reduces the load on the latch mechanism, while still permitting rotation so long as Z has a value greater than 1. The following example is based on a 75 mm diameter bore and piston:
Maximum axial force F 25,700 N,
Rod length adjustment range 4 mm,
Eccentric offset r 3.5 mm,
Eccentric rotational angle α+/−55 degrees,
Eccentric outside radius R 17 mm, and
Friction coefficient μ 0.05.
Reaction torque, Mr=Fr cos α=51.6 Nm
Friction torque, Mf=FμR=21.8 Nm
Net rotational torque Mn=Mr−Mf=29.8 Nm
  Z  =                    M        r                    M        f              =    2.37  
In this example rotation is possible and the net rotational torque Mn that is carried by the latch is reduced by the frictional torque Mf.
Except for the friction coefficient μ, the parameters affecting Z are geometric design parameters. The friction coefficient μ in contrast is only in part a function of the choice of design parameters including the eccentric 108 and rod small end 109 materials, contacting surface finishes, any coatings, and the lubricant formulation. It is also believed to be affected by hydrodynamic effects, particularly transient squeeze film lubrication driven by alternating connecting rod compressive and tensile forces acting on the journal bearing oil film between the connecting rod small end 109 and eccentric bushing 108. While not wishing to be bound by theory, it is believed that squeeze film lubrication is an important component of piston wrist pin lubrication, and is similarly important in determining the instantaneous friction coefficient μ between the eccentric and the rod, and that it is not adequately addressed in the prior art.
I. Elsion et al./Wear 261 (2006) 785-791 describes experimental investigation of piston wrist pin lubrication by rotationally oscillating a piston wrist pin within a journal that is clamped from opposite sides by an applied load. They report a mixed lubrication regime friction coefficient μ in the range of 0.03 to 0.06 for a steel pin and an aluminum journal, and show modest effects of experimental coatings and engineered surface finishes. Since the clamping load is constant rather than periodically reversing, these experiments do not provide information on transient squeeze film lubrication effects in which the bearing is fully hydrodynamic and there is no metal to metal contact. They do, however, provide information on the mixed lubrication friction coefficient when squeeze film lubrication is not occurring.
In squeeze film lubrication fluid forces momentarily separate two approaching solid surfaces in oil-flooded environments, forming a hydrodynamic bearing with a much lower friction coefficient than mixed lubrication. The hydrodynamic bearing support force FS for parallel configurations is given by:
      F    S    =      C    ⁢                            μ          l                ⁢                  L          4                ⁢        V                    h        3            
where μl is the lubricant viscosity at the operating temperature, L is the shortest flow path length from the center of the bearing area to the edge, V is the perpendicular velocity between the two surfaces, h is the separation between the surfaces, and C is a constant determined by the bearing geometry. It is believed that each time the rod alternates between compressive and tensile force, the journal oil film between the connecting rod small end 101 and eccentric bushing 102 forms a transient low friction hydrodynamic bearing. The time duration of this transient increases with oil viscosity μl and reduces with increased force F, resulting in variations with engine speed, load and temperature.
Transient low friction hydrodynamic bearing effects have the positive effect of facilitating eccentric rotation during rod length change, particularly during the rod length increases to raise compression wherein only relatively low inertial forces are available to provide the rotational torque. Conversely, transient low friction hydrodynamic bearing effects have the negative effect of increasing the torque load on the eccentric latch mechanism from cylinder pressure loading, particularly when maintaining maximum rod length during high compression operation.
In summary, prior art eccentric bushing variable compression connecting rods may not achieve reliable length shifts under all conditions, and have squeeze film bearing transients in the journal bearing between the connecting rod small end and eccentric bushing that are counterproductive in some operating modes. A need therefore exists for improvements that address these issues.